The present invention relates to a heat pumping installation, in particular with a refrigerator function.
Such installations are already used for the cold they produce and are applied for cooling purposes both in industrial processes (molding of plastics, manufacture of electronic components . . . ) and in the tertiary sector (distribution of foodstuffs, air-conditioning for computers . . . ) as well as for improving personal comfort (in cooling or air-conditioning systems in premises).
They have the advantage of avoiding the use of organic thermodynamic fluids in the compression-expansion cycle such as those belonging to the CFC family (chlorofluorocarbons), which have an adverse effect on global warming, or HCFCs (hydrochlorofluorocarbons) or HFCs (hydrofluorocarbons), which have a lesser but nonetheless not insignificant impact in terms of the greenhouse effect.
Their disadvantage, on the other hand, is the need to cope with very large volumes of vapor, particularly on a level with the compressor, which is one of the reasons why installations incorporating water vapor cycles have seen only very limited development to date.
Prototypes of such installations using water as the thermodynamic fluid and in cold-exchange and heat-exchange cycles have nevertheless already been built on an industrial scale. One of these, with a calorific output of some 2000 kW, used to cool extrusion machinery, uses an open production cycle to generate cold by evaporation, compression, condensation and discharge of water to the atmosphere, which constitutes a first disadvantage. It uses two independent steam compressors disposed face to face at the ends of a sealed, low-pressure enclosure, their suction inlets being arranged facing one another on either side of the evaporator, and these compressors, of the centrifuge type with flexible blades imparting to them a xe2x80x9cvariable geometryxe2x80x9d, being driven respectively by two electric motors, also of variable speed, outside the enclosure. Another disadvantage inherent in this type of installation resides in the fact that they require a large amount of space and carry a risk of air getting into the shaft ducts as well as heat losses as dissolved air gets into the installation via the open circuit of the condenser, which complicates the problem of degassing; on this issue, it should be pointed out that the non-condensable elements in this instance are drawn off at evaporation pressure, i.e. at low pressure. Furthermore, there is a susceptibility to relatively high xe2x80x9cnipsxe2x80x9d (differences between the exchange temperatures) on a level with the evaporator and the condenser.
Another, more compact prototype with a refrigeration output in the order of 800 kW is operated globally using the same thermodynamic cycle with water and also uses two separate compressors disposed inside the hermetically sealed enclosure along with their respective motors; although this approach solves the problem of sealing at the shaft ducts, the high peripheral velocity of the compressor wheels required to compress very large volumes of vapor, has meant designing them so that they use a blade structure made from carbon fibers, which imparts to them the necessary strength to withstand centrifugal forces but at the expense of service life, these wheels being very sensitive to erosion due to the impact of water droplets, incurring a risk that they will be driven at high speed at the suction end of the compressors.
Accordingly, the objective of this invention is to retain the advantages inherent in using water as a thermodynamic fluid but avoid the disadvantages of the techniques of the prior art in a heat pumping installation built to an industrial scale, the primary aim specifically being to produce cold but without ruling out the production of heat.
To this end, an installation proposed by the invention, of the general type outlined above, is characterized in that the refrigerant cycle uses a process of dynamic compression with two separate compression stages, linked to one another by at least one heat exchange zone (de-superheated and/or economizer) and contained in a steam confinement enclosure which is hermetically sealed and heat-insulated, and in that the wheels of these two sections are mounted directly on the opposite ends of the shaft of a common, sealed electric variable speed motor disposed inside said enclosure, between these stages.
Opting for a fully xe2x80x9cintegratedxe2x80x9d motor-compressor system of this type firstly makes for a more compact system and secondly overcomes the shaft sealing problem and, in a more economic manner, also resolves the tricky problem of designing a compressor capable of providing aerodynamic performance and advanced mechanical features whilst limiting the cost price of the installation. In particular, opting for a single electric motor to drive the two compression stages, each having one (in the case of compression by centrifuge, for example) or more (in the case of axial compression) compression wheel stages, and without the need to use speed multiplication stages, represents a decisive simplification in terms of structure. Furthermore, this design of confining the installation enables the compressor to be run without oil, thereby simplifying running and maintenance operations, whilst preventing fouling in the refrigerant fluid. It should be noted at this point that what are referred to as the xe2x80x9ccentrifugexe2x80x9d compression stages, which will be used by preference over axial compression stages, will comprise, in a conventional manner and for each of their constituent stages (of which there will be one or two in principle), a mobile wheel preceded by a suction convergent and followed by a static diffuser, either plain or provided with fins.
It should also be noted that the use of at least one vapor de-superheater between the two compression stages will prevent excessive temperatures from being reached, reduce the compression work of the second stage and help to improve the efficiency of the cycle, namely, will increase the ratio of refrigerant or calorific output to electrical energy needed to operate the installation, this efficiency possibly reaching a value of as much as 7 to 8, which is very satisfactory. This de-superheating after the first compression stage may be partially run by expansion-flash of the water coming from the condenser and returned to the evaporator, the expansion flash causing the water to be partially cooled without the need for any intermediate heat exchange surface, thereby constituting an economizer.
By preference, said electric motor will be a synchronous rotary motor with permanent magnets co-operating with a frequency controller, enabling the speed and hence the rotation speed of the compressor wheels to be varied to suit the vapor flows treated and enabling operation at partial load within the limits of the compressor""s aerodynamic stability. Opting for a motor of this type will ensure that there is a minimum of heat loss on a level with the rotor, which is an important factor given the poor heat exchanges achieved in an enclosure in which, when producing cold, the prevailing vapor pressure is very low. However, it would be conceivable to use other types of less expensive motors, for example asynchronous motors, with a device for eliminating heat losses.
The bearings for the shaft of said electric motor may be of any type suitable for the function they perform, for example ceramic roller bearings, or alternatively of the fluid or plain type, operated by water and having an anti-cavitation device, or even by oil and having a sealing device, or may be of the magnetic type, in which case it will be impossible for the refrigerant fluid to be contaminated by lubricant.
As a result of one feature of the invention, the shaft bearings for said motor are disposed to the side of the latter, the compressor wheels being mounted in an overhanging arrangement on the ends of the said shaft although the reverse layout is also possible: compressor wheels disposed between the motor and the bearings with no overhanging mounting.
Another feature of the installation resides in the fact that the two compression stages are disposed opposing one another on either side of the common electric drive motor, with their respective inlets (intakes) directed towards the ends of the confinement enclosure (contrary to the prior art described earlier), evaporation and de-superheating zones being provided between the ends of the enclosure and the inlet of the first and the inlet of the second compression stage respectively.
This layout provides compensation for the axial reactions due to the wheels, helps in obtaining greater compactness, particularly in terms of length, and facilitates connection to the external water circuits.
In situations where it would be necessary to increase the compression rate, particularly under certain climatic conditions, (when the external temperature is too high or there is too great a variance between the evaporation/condensation temperature), the two compression stages could also be linked to a third compression stage disposed inside the confinement enclosurexe2x80x94or placed in communication therewithxe2x80x94and provided as a booster disposed upstream or downstream of the compressor or alternatively between its two stages.
Advantageously, this booster will be driven by a hydraulic turbine driven on water borrowed in particular from the internal circuit, on a level with the evaporation or condensation stages but it could also be driven by a steam expansion turbine or an independent electric motor, optionally at a different speed from that of the compressor, which might even be at a standstill if there is a return to normal climatic conditions.
Advantageously and still with a view to reducing the cost price and easing the rotation loads, said booster or the compression stages may be provided as one or more compression wheels having a rotor with a rotating flange provided with radial flat vanes and optionally co-operating with static blading to pre-rotate the fluid.
The general layout of the installation may differ slightly depending on whether it has a booster or not: it will then be characterized, respectively, in that the condensation zone is located at the end of the confinement enclosure on the side of the suction inlet of the second compression stage or in that this condensation zone is located between the zone with de-superheating and this suction inlet of the second compression stage.